This invention relates to an improved differential pressure lubrication system for an eccentric rolling piston, sliding vane type of fluid compressor, as particularly used to compress refrigerant gases in refrigerators and air conditioners.
In conventional units of this type an electric motor and a rolling piston compressor driven thereby are mounted within a sealed pressure shell or casing. Refrigerant gases drawn in from an external accumulator or the like are compressed and discharged into the space within the shell, from which they flow to a condenser, evaporator or the like. A pool of lubricating oil is maintained within the shell and its surface is in direct contact with the high pressure discharge from the compressor. An oil flow path is established to properly lubricate the rolling and sliding friction members of the compressor such that the high pressure or supply end of the path is simply immersed in the pool of oil while the low pressure or return end is communicated with a suction passage of the compressor. The resultant differential pressure between the supply and return ends of the path establishes a steady flow of lubricating oil through the frictional members of the compressor. Such a relatively high differential pressure often produces an attendantly excessive flow of lubricating oil, however, which unduly loads the compressor, generates vibrations, results in an excessive amount of lubricating oil being entrained in the refrigerant fluid, etc.
In an effort to solve this "over-lubrication" problem, as disclosed in laid-open Japanese patent application No. 131393/83 and as shown in FIG. 1, the low pressure or return side of the oil supply passage is communicated with the compression chamber of the compressor in order to reduce the overall differential pressure to which the lubrication system is subjected. More specifically, by the action of an electric motor 2 mounted in a sealed shell 1, a crankshaft 3 is rotated to reduce the volume in a compression chamber 6 defined between a rolling piston 4 and a cylinder 5 to thereby compress refrigerant gases drawn in from an accumulator or the like, not shown. The compressed gases are released into the space 7 within the shell from which they are supplied to a condenser or the like via a discharge outlet 8. The lubricating oil 11 enters the compressor through a passage 9c formed in a side plate 9 and lubricates, in succession, bearing 9a adjacent end seal 12, eccentric 3a and bearing 10a in side plate 10. The oil then flows into the compression chamber 6 through a return passage 13 in the side plate 10, from which it is discharged together with the compressed gas into the space 7 within the shell and falls back into the supply pool. The bearings 9a, 10a have a relatively large clearance as exaggeratedly shown in FIG. 1 to establish a sufficient flow path for the oil, while the tolerance or clearance between the ends of the piston 4 and the side plates 9, 10 is relatively close to thereby effectively isolate the space 16 within the piston from the compression chamber 6. The necessary lubricating oil is supplied to the latter through the return passage 13.
Since the mean or average pressure in the compression chamber 6 lies between the suction pressure and the discharge pressure, with the latter being applied directly to the surface 11a of the oil pool, the differential pressure applied to the opposite ends of the oil flow path is thus considerably lower than in the more conventional arrangement described above, and this attendantly reduces the oil flow rate to thereby avoid such problems as undue loading, vibration, etc.
A disadvantage with the FIG. 1 approach is that the pressure at the bearing end 10b of the side plate 10 must be isolated from the discharge pressure within the space 7 in the shell. This requires a mechanical seal 14 which not only adds to the production cost, but also increases the mechanical loss due to friction and represents a further source of wear and deterioration. A further disadvantage is that the oil flow path includes successive restrictions represented by the bearing 9a, the clearance between the eccentric and the inner surface of the piston 4, and the bearing 10a, and even a partial blockage at any one of these points can result in overheating, seizure, and the destruction of the entire compressor unit.